Positive displacement rotary heat engine

ABSTRACT

The principle of compression in positive displacement, rotary engines working on basis of sinusoidal, cyclic compression and expansion is described and illustrated. The principle is adapted to a new, continuous flow, four stroke internal combustion engine with all compression-expansion chambers performing complete thermodynamic cycles for every revolution of the rotor, and the internal gas pressures acting on differential vane areas dynamically exposed during rotation. The operation of the engine is described, and estimated engine performances are shown for comparison with representative heat engines.

United States Patent [1 1 Isaksen 1 1 Oct. 2, 1973 1 1 POSITIVEDISPLACEMENT, ROTARY,

HEAT ENGINE [76] Inventor: Kjell Isaksen, 9737 SE. 41 St.,

Mercer Island, Wash.

[22] Filed: May 12, 1970 [21] Appl. No.: 36,605

[52] U.S. C1 418/218, 418/219, 123/835 [51] Int. Cl...... F0lc 1/00,F02b 53/08, F04c 17/00 [58] Field of Search 418/228, 229, 230,

[56] References Cited UNITED STATES PATENTS 2,154,457 4/1939 Knapp;418/218 969,353 9/1910 Evans... 418/217 X 1,464,408 8/1923 Collier..418/219 X 1,686,767 10/1928 Saxon 418/219 X 2,436,285 2/1948 Booth418/219 2,582,413 1/1952 Clark 418/153 2,724,369 11/1955 Barrett 418/211X 3,240,189 3/1966 Stumpfig.. 123/809 X 3,472,210 1 10/1969 Savoie418/191 X FOREIGN PATENTS OR APPLICATIONS 599,609 3/1948 Great Britain418/218 388,781 6/1965 Switzerland..... 893,197 1/1944 France 27,07111/1912 Great Britain 123/835 Primary Examiner-Carlton R. CroyleAssistant Examiner-John J. Vrablik [57] ABSTRACT The principle ofcompression in positive displacement, rotary engines working on basis ofsinusoidal, cyclic compression and expansion is described andillustrated. The principle is adapted to a new, continuous flow, fourstroke internal combustion engine with all compression-expansionchambers performing complete thermodynamic cycles for every revolutionof the rotor, and the internal gas pressures acting on differential vaneareas dynamically exposed during rotation. The operation of the engineis described, and estimated engine performances are shown for comparisonwith representative heat engines.

5 Claims, 24 Drawing Figures PATENTED UB1 2 75 SHEET 2 BF 9 i REFERENCElead.

trail.

REVOLVING A'NGLE- e b mm wDUmOP INVENTOR M Fig. 6

PAIENTED 013T 2I973 SHEET I [If 9 Fig.21A

IGNITION PATENTEUUET 21975 3.762.844

' SHEETSUF 9 INVENTOR zw wu PATENTEDBBT 2 3,762,844

SHEET 8 OF 9 64 FUEL INJECTlON 64 72 FUEL INJECTION SHEET 7 [IF 9PATENTEUUCT 2 1973 PATENTEB 2|973 9 1 10 cu.in.

Compress. Ratio: Swept Volume: Leak. Area 0.024 sq.in.

EST. COMPRESSION PERFORM.

SHEET 8 BF 9 ["Le m;

1 mDmwmmm CORR. RPM x 1000 Fig.19

.m U %c 2 2 1 mm V mm e .w S M 1T TI 5 E Rich mixture combust- CORR- RPMX 1000 Fig. 20

POSITIVE DISPLACEMENT, ROTARY, HEAT ENGlNlE CROSS-REFERENCES TO RELATEDAPPLICA- TlONS This patent application makes no cross-reference to anyprior, pending application.

BACKGROUND OF THE lNVENTlON This invention relates to a continuous flow,controlled compression ratio, positive displacement, rotary, heat enginewith radial vanes rotating with a centrally shafted, radial, wheel typerotor and forming a plurality of compression-expansion chambers inannuli between the rotor and the rotor housings, while moving in wipingcontact with sinusoidal walls of said housings, thereby exchangingtorque about the shaft of rotation when internal gas pressures act ondifferential vane areas dynamically exposed.

The positive displacement, rotary, heat engine as described in thisspecification is related in art to gas turbines, to reciprocating pistonengines and to a variety of rotary engines. The rotary movement and thecontinuous gas flow concept is inherited from gas turbines, where gaspressures also are acting on rotor blades to produce torque about acentral, rotating shaft. The positive displacement concept originated inthe reciprocating piston engines, where higher compression pressureswere easier attained than in the compressors of the gas turbines.

The double or multi-stage gas turbines are severely limited in utilityby the operating characteristic of their compressors with theirdifficulties in producing high pressures of compression over anyappreciable range of rotor speeds, and the difficulties involved inmatching teh compressors with the driving turbines. Difficultiesexperienced in cooling of the turbines have limited the operatingtemperatures in spite of introduction of costly turbine materials. Thegas turbines can therefore not be utilized to the full extent of theirthermodynamic posibilities. High costs and poor preformance have limitedthe general acceptance of the gas turbines.

In engines with reciprocating pistons acting on captive masses of gas,the heavy reciprocating parts severely limit the engine speeds andaccelerations. These engines are excessively heavy, requiring expensiveparts and regular maintenance. The combustion in these engines isgenerally very poor due to uneven distribution of fuels, and the exhaustgas emission of unburnt fuels and carbon monoxide is normally excessive.Four stroke operations are obtained through use of valve arrangements,which severely limit the flow and thereby the breathing capabilities ofthese engines.

Rotary engines generally fall in one of the following groups:

- derivates and improvements of the Wankel engine;

derivatives and improvements of the Cooley engine;

- derivatives and improvements of the two stroke,

eccentric, vane pump; and

other rotary engines.

In a Wankel engine a triangular piston rotates in an oscillatory pathinside a housing of elliptical shape. The movements of the rotatingpiston are controlled by an eccentric shaft geared to the piston.Compressionexpansion chambers are formed betwen the housing and thetriangular piston. This engine has until recently been a very poorperformer and is still inferior to a good reciprocating piston engine.The lack of performance has frequently been concealed by deceptiveinformation regarding the engine swept volume, which often reflectedonly one third of the real air breathing capacity. The engine has beenplagued with technical problems, some of which are inherent in the basicdesign. The engine lacks growth potential in its basic design and istherefore often found in compounded versions. The Wankel engine so farhas been unable to satisfy the high expectations advocated at itsconception, and it has so far made very limited impact on the enginemarket.

The original Cooley steam engine consisted of a two lobe, oval rotormoving in an oscillatory, rotational path inside a three clover leafshaped housing. Later models have used four lobes and five leaves. Thecompression-expansion chambers in this engine do not move with the rotoras in the Wankel engine, and a valve arrangement is essential. Thisengine is later adapted to the four stroke cycles of an internalcombustion engine in a variety of arrangements. This engine has alsobeen a very poor performer and has suffered from similar technicaldifficulties as the Wankel engine. No such engine is known to have beenmarketed.

In an eccentric vane pump the circular rotor mounted with radial vanesrotates inside an eccentrically located, circular housing. This engineis restricted to two stroke operations and is dependent on an externalsource for charging the compression-expansion chambers. The engineappears in a great variety of embodiments, none of which are known tohave been marketed.

A great array of geometrical engine designs of rotary types have beenpresented during the years. Most of these engines have never found anyacceptance. Rotary piston engines with cylinders and pistons rotatingaround a stationary shaft have been in use on early aircrafts, but thesewere later discontinued due to adverse operating characteristics.

Engines in use at this time are all seriously limited in theirapplication resulting from their operating characteristics or theexecution of their embodiments. A valid justification for pursuing a newdesign under the present conditions must be founded on the anticipationof gaining an advantage in one or more of the following main directions:

improvement in performance characteristics; reduction in weight;reduction in costs; and reduction in harmful air pollutants; over andabove known engines representing the state of the art. The inventiondescribed in this specification is expected to show significantimprovements in all of the stated directions.

SUMMARY The compression concept to be shown in this specification wasdeveloped to form the theoretical basis for a new type of heat engineintended to combine the best features of the conventional positivedisplacement, reciprocating piston engines and the continuous flow gasturbines.

In the engine described and shown, four stroke operating cycles can becompleted in a series of compression-expansion chambers during everycompleted rotation of the chambers in a single stage about the rotorshaft. All requirements for valve arrangements as in the reciprocatingpiston engines and all requirements for separation of compression andexpansion into different stages as in the gas turbines are eliminated.Flow of fluids into the engine through the intake ports and out of theexhaust ports are continuous.

Fluid compression is achieved by confining a volume of fluid within anenclosed chamber, which is then compressed in volume during a rotarymovement of the rotor with respect to the rotor housing. The compressionis controlled by relieving the fluid in the compressing chamber into asecond, expanding chamber, thereby creating a flow of fluids between thetwo chambers. In the case of an internal combustion engine, combustionwill take place during this cross-over flow, injecting burning gasesinto the expanding chamber. The pressures in the compressing chamberwill always exceed those in the expanding chamber in order to maintainfluid flow. The exposed differential work area will, however, soonbecome positive in value and positive work will result.

The internal fluid pressures act on differentially exposed vane areasduring the chamber rotation about the rotor shaft. The exposed areas arenegative during compression and positive during expansion.

Fluid induction, exhaust and ignition, as applicable, form integratedfunctions in the thermodynamic cycles.

The illustrated embodiment of an internal combustion engine is selectedfrom a number of possible embodiments all belonging to the same familyand described by the same basic operating principles. The engine sizeand the compression ratio was selected especially to compare withconventional positive displacement, reciprocating piston engines incommon use.

The performance of the shown engine has been analyzed in a conservativemanner and is illustrated in the figures for comparative purposes. Theshown performance curves are limited to the shown embodiment and thestated engine displacement. Since a variety of embodiments is possiblewithin the operating prinicples, special performance characteristics mayeasily be developed for special requirements. No statement in thisspecification shall thus be construed to mean that an optimum engineembodiment is shown.

The invention is limited in scope by the described prinicple(s) ofcompression and by the stated claims essentially as described in thisspecification and supported by the attachedillustrations.

The shown embodiment of the internal combustion engine has been designedwith the object of securing the following relative advantages:

- a high power/weight ratio;

a high. specific power output;

an extended operating speed range;

an improved operating economy;

a rotor assembly in complete balance;

a compact packaging;

a simplified mechanical operation;

a simplified embodiment;

a simplified accessory requirement;

a full volume, forward flow combustion; a controlled heat transfer rate;

reduced port losses;

complete rotor burst containment capability; reduced toolingrequirement; reduced material costs; and

with extensive growth potential all in comparison with the known stateof the art in positive displacement, reciprocating piston enginetechnology as far as comparison is possible.

For the purpose of easing comparison with the reciprocating pistonengines in use, the illustrated internal combustion engine embodimentperformances will be based on rich mixture combustion with fuelintroduced in conventional manner through a carburettor or injected intothe intake manifold through a metering systern. This method ofcombustion will be referred to as full volume combustion as compared topartial volume or partial flow combustions as described later .in thisspecification.

BRIEF DESCRIPTION OF THE DRAWINGS The following illustrations areoffered in support of this specification to describe the basiccompression principlc(s) involved and the mechanical embodiment with ananalysis of the engine performances during operations:

FIG. 1 illustrates the basic prinicple of compression in the simplesttwo-stroke concept;

FIG. 2 illustrates the basic principle of compression as applied to thebasic four-stroke concept;

FIG. 3 illustrates the relationship between the compression-expansionchamber sector angle and the compression ratio for the cases of FIG. 1and FIG. 2

FIG. 4 illustrates the boundaries of a compression chamber in elevationand defines the reference line;

FIG. 5 illustrates the volume variations of a compression-expansionchamber as the rotor revolves about the rotor shaft for the caseidentified in FIG. 3;

FIG. 6 illustrates the differentially exposed vane areas as thecompression-expansion chamber of FIG. 5 and the rotor revolves about therotor shaft;

FIG. 7 illustrates the four-stroke, internal combustion engineembodiment in exterior side elevation;

FIG. 8 illustrates the same engine embodiment in a transverse sectionalong line AA of FIG. 7 and FIG.

FIG. 9 illustrates the engine embodiment in exterior side elevationviewed from the opposite side to FIG. 7;

FIG. 10 illustrates one of the two profile housings in exteriorelevation and with the face plate removed;

FIG. 11 illustrates a section of FIG. 10 and FIG. 12-

along the line 8-3;

FIG. 12 illustrates the profile housing in interior elevation;

FIG. 13 illustrates one of the two identical rotor halves as viewed fromthe oposing rotor half;

FIG. 14 illustrates a section of the same rotor half along line C-C ofFIGS. 13 and 15;

FIG. 15 illustrates the above rotor half as viewed from the matingprofile housing;

FIG. 16 illustrates details of a rotor vane;

FIG. 17 illustrates the exposed torque areas of two adjacent rotor vanesenclosing a compressionexpansion chamber;

FIG. 18 illustrates the maximum movements of a rotor vane in the rotorand locates the vane center of gravity;

FIG. 19 illustrates the mean pressure of compression of the dividedcompression-expansion chamber as its centerline passes the top deadcenter; and

FIG. illustrates the estimated net torque and the brake horse power ofthe engine as related to rotor speeds.

FIG. 21A shows a schematic of a circumferential view through thecompression-expansion chambers while in suction and compression modes ofthe cyclic thermodynamic four stroke operation.

FIG. 21B shows a schematic of a circumferential view through thecompression-expansion chambers while in ignition, expansion and exhaustmodes of the cyclic thermodynamic four stroke operation.

FIG. 22 shows a schematic of a circumferential view through thecompression-expansion chambers while in compression, ignition andexpansion modes of a four stroke thermodynamic cycle operation involvingfuel injection into a moving enclosed airstream under compression and/orexpansion followed by auto-ignition of the mixture.

FIG. 23 shows a schematic of a circumferential view through thecompression-expansion chambers while in compression, ignition andexpansion modes of a four stroke thermodynamic cycle operation involvingfuel injection into a moving enclosed airstream under compression and/orexpansion followed by ignition of the mixture by and external source.

DESCRIPTION OF THE PREFERRED EMBODIMENT The invention and the preferredembodiment is based on the following principle of compression:

In FIG. 1 a sine curve is shown with its axis displaced by a fullamplitude or radius to become a tangent to the minimum point on thecurve. The area confined by the sine curve, the displaced axis and adefined angular span described as a constant fraction of thecircumference or the revolving angle, varies with the location of thisangular span around the circumference as shown in two positionsidentified by shaded boundaries in FIG. 1 and FIG. 2. The ratio of themaximum to the minimum areas confined may be defined as the areacompression ratio. The compression ratio is controlled by the magnitudeof the fraction of the angular span to the circumference or therevolving angle and is independent of the sine curve amplitude or radius(r) or by the magnitude of the radius of rotation as long as thesequantities remain constant around the circumference. Since the areacompression ratio is independent of the radius of rotation, the areacompression ratio must be identical to the volume compression ratio forall cases also confined between any two radii of rotation.

FIG. 2 shows a two cycle sine curve confining compression areas in thesame manner as shown in FIG. 1.

In FIG. 3 the compression ratios of both single and double cycle sinecurve operations are shown as functions of the angular span or sectorangle. It is seen from this figure that decreasing angular span orsector angle results in increasing compression ratio.

The illustrated principle of compression is different from the principleof compression employed in rotary pumps by the fact that means ofcontrolling the compression ratio is available, while in pumps thecompression ratio invariably approaches infinity as the minimum volumeapproaches zero. The illustrated principle of controlled compressionratio has no meaning in rotary engines, however, unless the dividedminimum volumes are interconnected by a flow passage. Overcompression ofthe fluids in the trailing, compressing part of the chamber is therebyrelieved by flow into the leading, expanding part of the chamber, and acompromise compression ratio and thereby a fluid pressure level isreached. When no such passage is available, a compression ratio ofinfinity is approached as the part of the chamber on the trailing sideof the compression peak approaches zero. The presence of such a flowpassage, regardless of form, tends to reduce the compression ratiosbelow the theoretical values shown in FIG.

Rotary engines employing single cycle sine curve contours are limited totwo-stroke thermodynamic operations, while double sine curve contourswill accept both two and four-stroke thermodynamic operationalprocesses. From the magnitudes of the compression ratios for the singleand the double cycle sine curves as shown in FIG. 3, it is concludedthat the shown principle of compression for the stated curves may beapplied in practical thermodynamic process operations in the mannershown in FIGS. 1 and 2. It can be shown that triple and quadra cyclesine curves require increasingly more sector divisions or smaller sectorangles to secure practical compression ratios, thus reducing theprospects for adaption in simple embodiments.

From the possible solutions shown in FIG. 3, the special case marked 0on the double cycle sine curve re lationship between sector angles andcompression ratios is selected for embodiment as a four stroke, positivedisplacement, internal flow combustion, heat engine.

The compression-expansion chamber is illustrated in FIG. 4 as shown inelevation. The reference line is drawn on this figure for the purpose ofpositive identification of location of the chamber in laterillustrations. The single compression-expansion chamber in the fig ureis enclosed by two concentric walls in axial direction at the two radiiof revolution R, and R by a wall in the plane of the paper illustratingthe displaced axis, by a concentric double cycle sinusoidal plane facingthe paper and by two sector walls of variable areas in an axial plane tocomplete the enclosure of the chamber. By introducing a relative motionbetween the sinusoidal wall and the remaining walls in a circulardirection, the two sector walls will vary in exposed areas in accordancewith the respective distances between the sinusoidal plane and thedisplaced axis plane. Any pressure arising inside thecompression-expansion chamber enclosed by the stated walls will createforces on all the enclosing walls. Selecting the sinusoidal wall for astator and allowing the remaining walls to rotate about the centralaxis, the internal chamber pressures create torque about this axis inaccordance with the exposed areas of the sector walls and their armswith respect to the axis. When the exposed sector wall A is greater thanthe exposed wall A a positive torque is created about the axis ofrotation. When the magnitudes are reversed, the torque about the axis isnegative, and when the two areas are equal, the chamber is at a deadcenter.

FIGS. 5 and 6 illustrate the magnitude of the compression-expansionchamber volume and the differential vane or torque areas in percents asfunctions of revolving angular locations of any one of the six chambersabout the rotor shaft with respect to their reference lines. Bycross-relating the two figures and applying the above sign convention tothe differentially exposed sector wall areas defined as torque area, (A

A it is easily seen that compression requires work, while expansiongives up work.

The embodiment of the stated internal combustion engine is shown inFIGS. 7 through 18 attached to this specification.

Referring to the figures, FIGS. 7, 8 and 9, the engine power unitcomprises a rotor assembly splined and shrunk onto a shaft 31 and freeto rotate in two double bearings 32 and 33. The bearings 32 and 33 arehoused in two opposing profile houses 34 and 35. The two profile houses34 and 35 enclose the rotor assembly 30 and are kept in a fixedrelationship by means of an oil col lector ring 36 serving as a spacer.The exteriors of the two profile houses 34 and 35 are covered by twoface plates 37 and 38 to enclose the coolant passages 39 and 40 and thelubrication annuli 41 and 42 in the profile houses 34 and 35. The faceplates 37 and 38 and the profile houses 34 and 35 are penetrated by twoinlet ports 43 and 44, two exhaust ports 45 and 46 and two ignitor plugholes 47 and 48. The face plates 37 and 38 are also penetrated by twocoolant inlet holes 49 and 50, two coolant outlet holes 51 and 52, twolubrication inlet holes 53 and 54 and two lubrication overflow holes 55and 56.

The power unit is held together and mounted by fasteners through theholes 57 and 58. The face plates 37 and 38 are also fastened to theprofile houses 34 and 35 through the attachment holes 59 and 60 and also61 and 62. Provisions are made to attach an oil sump at the surface 63.

Six vanes 64 penetrate the rotor discs 65 and 66 through six radialslots 67 as shown in FIGS. 8, 13, 15, 16, 17 and 18. The vanes 64 aresupported by pivots 68 at the shaft end of the rotor assembly 30 in therotor hubs 70 and 71. The rotor vanes 64 move in soft or elasticallyloaded wiping contact with the sinusoidal walls 72 and 73 of the profilehouses 34 and 35 through a set of soft loaded sealing strips 74 at theedges of the rotor vanes 64.

Twelve compression-expansion chambers, six on each side of the rotorassembly 30, are formed between the two opposing profile houses 34 and35, the rotor discs 65 and 66, the rotor vanes 64, the rotor hubs 70 and71 and the outer rotor rims 75 and 76. The sinusoidal walls 72 and 73 ofthe profile houses 34 and 35 are off-set by half a sinusoidal wavelength to allow room for the rotor vanes 64.

The inlet ports 43 and 44 and the exhaust ports 45 and 46 of the profilehouses 34 and 35 are inclined to reduce the turning angle of theincoming and exhausting gases as seen from FIGS. 7, 9 and 10. Both theinlet ports 43 and 44 and the exhaust ports 45 and 46 are so locatedthat they provide the correct timing for the thermodynamic cycles asshown in FIG. 12. The four engine strokes shown in this figure arereferred to the reference line previously identified. Sector I-Iindicate the suction stroke, sector IIII the compression stroke, sectorIII-III the power expansion stroke and sector IV-IV the exhaust stroke.Port overlap is indicated by sector VV.

Passages 77, shown in FIG. 15, are cut in the rotor discs 65 and 66 toprovide connection between the two divided parts of thecompression-expansion chambers over the compression top in addition tothe flow passages created by the running clearances. These passages 77must be sized to prevent any significant pressure backup in the trailingpart of the compressionexpansion chamber and to secure the correct flowve-,

locities to the compressed gas during the passage. Excessive sizing ofthe flow passage 77 will reduce the compression ratio needlessly. Theigniter plug holes 47 and 48 are positioned near the top dead center tocoincide with the flow passages 77. ignition of the flowing gas streamcan in this manner be timed well in advance of the compression pressurepeak.

Gas seals are provided at the rotor vane edges 74, in recesses '78 inthe rotor disc slots 67 and in the groves 79 and 80 at the edges of thesinusoidal walls 72 and 73 of the profile houses 34 and 35. These sealsare introduced to improve the low rotor speed operating characteristics.The seals are designed and loaded to provide maximum sealing for minimumfriction losses. Details of the pneumatically loaded seals are notshown.

Two separate cooling systems are used due to the differences in coolingrequirements and the need for return of the lubricating oils.

The sinusoidal walls 72 and 73 of the profile houses 34 and 35 arecooled by liquid coolant entering the cooling passages 39 and 40, asshown in FIGS. 7, 9 and 10, through holes 49 and 50, counter-flowing theinternal gas flow direction in the compression-expansion chambers andleaving through the holes 51 and 52.

Cooling of the rotor assembly 30 is afforded by means of lubricatingoils entering the oil annuli 41 and 42 at the holes 53 and 54. The oilthen enters the main shaft 31 through the double bearings 32 and 33 andpasses into the rotor assembly 30 at the pivot gallery 81, as shown inFIG. 13, for lubrication of the pivot bushings and the pivot pins 69 and68. From the pivot gallery 81 the oil distributes into radial coolingpassages 82 between the two halves of the rotor assembly 30. The coolingpassages 82 are throttled at the radial exits 83 to provide bettercontact between the oil and the rotor disc walls 65' and 66 during heattransfer and to reduce the cooling flow rate at higher rotor speeds. Theejected oil is then collected in the collector ring 36 and is drainedback into the oil sump after some splashback against the outer rotorrims 75 and 76. Oil cooling takes place in the oil collector ring 36 andin the oil sump.

MODES or OPERATION The engine operations are shown in FIGS. 21A, 21B, 22and 23 and are described as follows:

When a leading vane 64 of a compression-expansion chamber passes thecompressing peak of the sinusoidal wall 72 or 73 between the exhaust 45or 46 and the intake ports 43 or 44, a suction is created in theexpanding part of the chamber. Air or air/fuel mixture is drawn into theexpanding chamber through the intake port at moderate velocities. Assoon as the chamber volume has reached its maximum and the trailing vane64 of the chamber passes the end of the intake port, 43 or 44, thecompression stroke commences.

As soon as during continued rotor movement the leading vane 64 of thecompression-expansion chamber reaches the compression peak at theigniter plug location in hole 47, the compression-expansion chamberdivides into two parts. The leading part of the chamber expands involume at an initially slow rate, while the trailing part compresses ata higher rate. The pressure differential thereby created between the twoparts of the chamber causes a flow of fluid from the trailing and intothe leading part of the chamber through the flow passage 77 at avelocity higher than the rotor speed. At a timed position the flow isignited, and a flame front propagates into the combustible mixtures atincreasing velocities relative to the advancing local flow velocity.After a relatively slow start the pressures in the compression-expansionchamber build up in both the divided parts. The flow between the twoparts of the chamber continues during the pressure increase, while theflow velocities and the flow densities adjust to the new conditions. Theflow combustion continues until either fuel or air supply is exhausted.

During the power-expansion stroke the leading vane 64 increases inexposed area until a maximum is reached. The trailing vane 64 areaexposed during this stroke reduces in magnitude until it reches theminimum size as it passes the sinusoidal wall compression peak at theigniter. The differential vane areas are therefore positive. Thepressure in the compressionexpansion chamber reaches its maximum valueduring combustion a short time after the minimum chamber volume haspassed and the volume is increasing and then reduces sharply during thechamber expansion as the positive value of the differentially exposedvane or torque area increases.

As soon as the leading vane 64 of the compressionexpansion chamberreaches the exhaust port 45 or 46, the remaining chamber pressurecollapses as the exhaust gases expand into atmosphere. The collapse rateand thereby the exhaust noise can to some degree be controlled by thelocation and shape of this exhaust port.

The scavenging of the compression-expansion chamber is assisted by thechamber compression following the expansion stroke. The port overlap inthis engine has a difierent meaning than in a reciprocating pistonengine. The port overlap in this engine improves the scavenging, butsince the chamber is divided into two parts with only a small passage77, the inertia effect of the exhaust gases to improve the inflow of newgases through the intake port 43 or 44 is of reduced value. Thecontinuous inflow through the intake ports 43 or 44 will in any casesecure the best possible filling of the chambers.

The process cycle then repeats itself.

Twelve identical process cycles are executed for every revolution of theengine rotor, six on each side of the rotor assembly. The shownembodiment will execute a power stroke for every 30 movement of therotor. The power strokes are thereby overlapping, and the powerfluctuations are consequently reduced to a minimum. The moment ofinertia created by the rotor itself is therefore adequate to secure aneven, continuous rotor rotation, and no requirement for any flywheelexists. A reciprocating piston engine with four stroke operation willrequire 24 cylinders to execute the same number of power strokes pershaft revolution.

The shown embodiment indicate that fuel/air mixing takes place beforethe combustible mixtures enter the compression-expansion chambers. Inthis manner the fuel/air mixtures fill the entire compression-expansionchamber before the compression, and an essentially full volumecombustion will take place. Care must be taken in sizing the flowpassage '77 to avoid overcompression of the combustible mixtures in thetrailing part of the compression-expansion chamber at higher rotorspeeds, which otherwise may cause detonations. The

flow type combustion is a considerable improvement over combustion byflame front propagation in a relatively stagnant combustible mixture, asis the case in reciprocating piston engines. Identical distribution offuel/air mixtures to all compression-expansion chambers makes leanmixture combustion possible with improved fuel economy and reducedamounts of unused fuels and carbon monoxide in the exhaust gases with noapparent loss in power.

A different method of combustion may be utilized when fuel is injecteddirectly into the compressionexpansion chamber during the compressionand after a part of the chamber has passed the compression peak and isunder expansion on the leading side. The chamber will then receive fuelinjection into the trailing part of the chamber only. When thiscombustible mixture is forced through the flow passage 77 and ignited, aburning jet of fuel/air will be injected into the leading part of thechamber containing air only. Excess fuels injected into the trailing,compressing part of the chamber will be burnt off in the excess air inthe expanding part. Extremely low levels of carbon monoxide and unburntfuels will result from this type of combustion. This method waspreviously referred to in this specification as the partial volumecombustion.

A third method of combustion is also possible in this engine. Thismethod was previously referred to as the partial flow combustion method.By special forming of the the flow passage 77 into different paths it ispossible to divide the flow passing over the compression peak. Injectingfuel into the air passing through some of these passages for combustion,while letting the remaining flow paths by-pass the combustion, will makea partial flow combustion possible with a down-stream mixing ofcombusted and non-combustible gases with a possible burn-off of excessfuels during expansion. This method of combustion has the potential forcomplete combustion control leaving a minimum of unburnt fuels andcarbon monoxide in the emitted exhaust gases. Combustion in excessiveair as described in the preceeding combustion methods, will result ingreatly improved fuel economy.

As seen from the above description the engine will accept a carburettorsystem, inlet port fuel injection and, with some modification to theshown embodiment, fuel injection directly into the compression-expansionchamber in at least the two previously described manners. Since theengine operates on continuous flow through the inlet ports, identicalcharges will reach every chamber on each side of the rotor. The collingsystem of the engine will also permit higher combustion temperaturesthan usual in reciprocating piston engines without any decay in thechamber wall metallurgical properties. it is therefore possible tooperate this engine much closer to the best power fuel/air ratio than inprevious positive displacement engines. A fully lean mixture operationis proposed with fuel enrichment towards best power mixtures at higherrotor speeds and torque requirements, instead of the present richmixture operation with further fuel enrichment under more demandingoperations. Two point fuel injection can be provided for best possiblefuel control at moderate costs.

The shown engine embodiment will accept a conventional type circuitbreaker in the ignition system geared to give two series of six sparksper rotor revolution. A more advanced electronic system may be requiredto enable engine utility at higher rotor speeds. No distributor isrequired for this system. Two ignition coils may instead be used inconjunction with the spark generator and the sparking plugs.

Transmission of torque from the vanes 64 follow the load path throughthe pivots 68 and bushings 69 to the rotor assembly 30 and the shaft 31.The gas seals recessed in groove 78 in the disc slots 67 are elasticallyloaded and do not transmit any appreciable load directly to the rotor30. The torque loads experienced by the pivot pins 68 are thussuperimposed on the radial loads caused by rotation. Since the vanes 64are soquentially loaded on either side, the two sides of the vane pivotpin 68 will be subjected to an alternating load pattern.

The vanes 64 are made with hollow core to maintain a center of gravityclose to the pivot 68 as shown in FIG. 18 This close center of gravityposition reduces the vane loads on the sinusoidal walls 72 and 73 andimprove the rotor speed range.

The working temperatures of the rotor assembly 30 and the sinusoidalwalls 72 and 73 must be controlled to levels below their metallurgicallimits and also below the critical temperature for ignition of thecombustible mixtures prematurely. The wall temperatures are controlledby creating a temperature gradient over the walls by means of heattransfer to a flowing heat transfer fluid or coolant on the outside ofthe wall. The cyclic heat transfer losses are greatest at lower rotorspeed when the combustion temperatures are lower and the time exposuresare longer during each cycle. At higher rotor speeds the combustiontemperatures stabilize, and the heat transfer rate at shorter timeexposure becomes constant in each operating sector reducing therequirement for increasing coolant flow rates. Cooling of the rotorassembly 30 is afforded by returning the lubricating oil through thecore of the rotor discs 65 and 66 on its return to the oil sump. Thiscooling requirement is minor due to the rotor exposure to the coolerinflowing gases during its rotation. The cooling requirements for thesinusoidal walls 72 and 73 vary around the circumference. The greatestcooling requirement exist in the combustion sector of the wall, which iscontinuously exposed to the high temperatures of combustion. The coolingfluid is therefore pumped in near the exhaust ports 45 and 46 and outnear the intake ports 43 and 44 to prevent warping of the structure andreduce the wall temperatures. The rotor vanes 64 are cooled byconducting heat into the adjoining cooled members of the structure.

Fuel requirements for the engine varies with the method adapted forintroduction of the fuel and the compression ratio selected. Normaloctane value automotive gasoline may be used in conjunction with acarburettor for the engine in shown embodiment. The octane value of thefuels may be substantially reduced when fuel is injected directly intothe compressionexpansion chamber in the combustion sector in a mannerdesigned to prevent compression ignition of the combustible mixtures outof sequence. Lower octane fuels emit in many cases less harmfulpollutants and are more completely consumed during combustion.

FIG. 19 shows the mean pressure of compression capability of thedescribed engine in a two liter displace ment version and operatingunder adverse leakage conditions. It is seen that as the rotor speedincreases, the effect of leakage on the pressures of compressionreduces. The shown pressures are mean values of the pressures existingin the leading and trailing part of the compression-expansion chamber.The pressure differential between the two chambers may not exceed twopercent of the differential value for the shown compression ratio.

In FIG. 20 an estimate is shown of the torque and power performance ofthe described engine in a two liter version as function of rotor speeds.Since the estimate was based on adverse leakage conditions, the torquedeveloped by the engine is relatively low at lower rotor speeds andpeaks relatively late in the speed range. Allowances have been made inboth curves for operation of essential accessories.

The specific power output of the shown two liter displacement version ofthe engine is approximately 2.5 B.H.P. per cubic inch displacement ascompared with approximately 1.2 for a normal high performance automotiveengine of similar displacement and even less for a Wankel engine. Theestimated power/weight ratio depends heavily on the choice of materielsfor the structure and also on accessories. A power/weight ratio of 2.9B.H.P. per lb. weight seems reasonable for the power unit alone in thetwo liter displacement version, and a lower value with all accessoriesand in running condition. Comparable values for positive displacementengines in running condition do not exceed I.0 B.I-I.P. per lb. weight.

ALTERNATIVES The principle of compression as described in thisspecification is adaptable to a great variety of applications incompressible fluid machinery. The comparative advantage, however, seemsto be connected to machinery using internal compression followed byexpansion and preferably with addition of heat or other pressureincreasing process occuring in conjunction with the two strokes.

FIG. 3 clearly shows the relationship between the division intocompression-expansion chambers and the compression ratio. A departurefrom the true sine or cosine curve form will change this relationship.The expression sinusoidal wall in this specification is used in themathematical sense within a band of reasonable tolerances to describe awall of sinusoidal character regardless of the method used in ariving atthis curve. This invention as expressed and shown covers all theindicated compression possibilities along the the curves in FIG. 3 forthe type of structure shown in this specification.

The introduction of pivoted vanes is closely connected with requirementfor high performance. Varies sliding in slots will perform adequatelyfor many purposes where high performance is not required and the wearrate is not of great importance.

The double acting engine shown in this embodiment improves the packagingand the performance. The two sides operate thermodynamicallyindependent. Alternative methods for securing continuous contact betweenthe vanes and the sinusoidal wall may be used as application requires.

When the compression pressures become adequately high, compressionignition may be adapted instead of electrical ignition. Fuel injectiondirectly into the compression-expansion chamber then becomes imperative.

The engine is growth limited by the relative Mach. no. between the rotorand the entering gas at the intake ports 43 and 44 and the ability tocool the sinusoidal walls of the profile houses 34 and 35 at higherengine speed operations. The shown two liter displacement engineoperates so far from these critical limits that considerable growth maybe possible before any reduction in rotational speeds becomes essential.

With reference to the preceeding specification supported byillustrations, I hereby respectfully request patent protection for thefollowing claims:

1. A positive displacement, rotary, heat engine, comprising:

a wheel type, radial rotor 30 extending into an expanded rotor rim 75,securely attached to a shaft 31 at the center of the rotor hub 70 andmounted for rotation;

said rotor 30 facing a housing 34 extending radially outwards from saidshaft 31 in spaced relation to said rotor 30;

said rotor 30 in conjunction with said housing 34 enclosing an annularchamber;

the inner surface of said housing 34 facing said annular chamber beingcontoured as a sinusoidal wall 72 and converging towards focal points inthe proximity of said rotor shaft 31;

said chamber having means of flow passage 77 over the sinusoidal peaksindented into the face of the rotor disc 65;

said rotor 30 having a plurality of radial slots 67 disposed atequi-angular location, the exact number of said slots 67 beingdetermined by the compression ratio requirement on basis of describedmathematical principles; I

-- a rotor vane 64 free to move in axial direction in each slot 67 andhaving means to maintain edgewise wiping contact with said saidsinusoidal wall 72;

said rotor 30, said sinusoidal wall 72 of said housing means 34 and saidrotor vanes 64 defining a plurality of sinusoidal expanding andcontracting annular sector chambers during rotation of said rotor 30relative to said housing means 34;

- said rotor vanes 64 moving in wiping contact with said sinusoidal wall72 being differently and dynamically exposed to fluid pressures in saidsector chambers to negotiate torque about said shaft 31 of revolutionduring rotational movement of said rotor 30 relative to said housingmeans 34; I

-- said housing means 34 having intake 43 and exhaust 45 ports locatedfor direct periodic communication into said sector chambers andextending circumferentially over predetermined sectors to control theprocess timing of predefined thermodynamic cycles as said chambers moverotarily relative to said housing means 34;

- said rotor 30 having forced radial fluid flow means 82 for internalcooling; and

said housing means 34 having forced fluid flow means for cooling saidsinusoidal wall 72.

2. In an engine as described in 1.:

said rotor vanes 64 being pivotally mounted 68 in the hub 70 of saidrotor 30 for transmittal of substantial loads through said pivots 68;and

- having elastically loaded sealing means in said rotor slots 67 and atsaid rotor vane 64 wiping edges 74.

3. in engines as described in 1.:

- means located in said sinusoidal wall 72 in said housing means 34 forsequential injection of fuels into flowing reacting agents in saidsector chambers.

4. ln engines as described in 1.:

igniter means 47 located in said sinusoidal wall 72 of said housingmeans 34 and operable for sequential ignition of compressed combustiblemixtures flowing through passages 77 in said sector chambers.

5. In engines as described in 1.:

means located in said sinusoidal wall 72 in said housing means 34 forsequential injection of fuels into flowing reacting agents in saidsector chambers: and

igniter means 47 located in said sinusoidal wall 72 of said housingmeans 34 and operable for sequential ignition of compressed said fuelsand reacting agent mixtures flowing through said passages 77 in saidsector chambers.

1. A positive displacement, rotary, heat engine, comprising: - a wheeltype, radial rotor 30 extending into an expanded rotor rim 75, securelyattached to a shaft 31 at the center of the rotor hub 70 and mounted forrotation; - said rotor 30 facing a housing 34 extending radiallyoutwards from said shaft 31 in spaced relation to said rotor 30; - saidrotor 30 in conjunction with said housing 34 enclosing an annularchamber; - the inner surface of said housing 34 facing said annularchamber being contoured as a sinusoidal wall 72 and converging towardsfocal points in the proximity of said rotor shaft 31; - said chamberhaving means of flow passage 77 over the sinusoidal peaks indented intothe face of the rotor disc 65; - said rotor 30 having a plurality ofradial slots 67 disposed at equi-angular location, the exact number ofsaid slots 67 being determined by the compression ratio requirement onbasis of described mathematical principles; - a rotor vane 64 free tomove in axial direction in each slot 67 and having means to maintainedgewise wiping contact with said said sinusoidal wall 72; - said rotor30, said sinusoidal wall 72 of said housing means 34 and said rotorvanes 64 defining a plurality of sinusoidal expanding and contractingannular sector chambers during rotation of said rotor 30 relative tosaid housing means 34; - said rotor vanes 64 moving in wiping contactwith said sinusoidal wall 72 being differently and dynamically exposedto fluid pressures in said sector chambers to negotiate torque aboutsaid shaft 31 of revolution during rotational movement of said rotor 30relative to said housing means 34; - said housing means 34 having intake43 and exhaust 45 ports located for direct periodic communication intosaid sector chambers and extending circumferentially over predeterminedsectors to control the process timing of predefined thermodynamic cyclesas said chambers Move rotarily relative to said housing means 34; - saidrotor 30 having forced radial fluid flow means 82 for internal cooling;and - said housing means 34 having forced fluid flow means for coolingsaid sinusoidal wall
 72. 2. In an engine as described in 1.: - saidrotor vanes 64 being pivotally mounted 68 in the hub 70 of said rotor 30for transmittal of substantial loads through said pivots 68; and -having elastically loaded sealing means in said rotor slots 67 and atsaid rotor vane 64 wiping edges
 74. 3. In engines as described in 1.: -means located in said sinusoidal wall 72 in said housing means 34 forsequential injection of fuels into flowing reacting agents in saidsector chambers.
 4. In engines as described in 1.: - igniter means 47located in said sinusoidal wall 72 of said housing means 34 and operablefor sequential ignition of compressed combustible mixtures flowingthrough passages 77 in said sector chambers.
 5. In engines as describedin 1.: - means located in said sinusoidal wall 72 in said housing means34 for sequential injection of fuels into flowing reacting agents insaid sector chambers: and - igniter means 47 located in said sinusoidalwall 72 of said housing means 34 and operable for sequential ignition ofcompressed said fuels and reacting agent mixtures flowing through saidpassages 77 in said sector chambers.